Hydrostatic bearings for the swash plate of a barrel-cylinder hydraulic pump or motor

ABSTRACT

This hydrostatic bearing for the swash plate of a hydraulic high-pressure and cylinder barrel machine is formed on a face of a bearing plate which registers with one face of a swash plate, and comprises a main bearing consisting of a set of two arcuate grooves surrounded by sealing surfaces and disposed in proper phase relationship to ports formed in the fluid distributor plate, and another set of grooves in the other face of said swash plate, identical in number to that of distributor orifices formed at the end of the barrel cylinders, said bearing further comprising at least three auxiliary hydrostatic bearings consisting of cavities formed in the other face of said bearing plate, said last-named cavities being concentric to the grooves of said main bearing.

[ Aug. 6, 1974 3,040,672 6/1972 Foersten et 91/487 3,073,253 1/1963Schollhammer........... 91/485 3,635,126 1/1972Engel.........,...,.......................91/486 PrimaryExaminer-William L. .Freeh Attorney, Agent, or Firm-Stevens, Davis,Millcr & Mosher [57] ABSTRACT This hydrostatic bearing for the swashplate of a hydraulic high-pressure and cylinder barrel machine is formedon a face of a bearing plate which registers with one face of a swashplate, and comprises a main bearing consisting of a set of two arcuategrooves surrounded by sealing surfaces and disposed in proper phaserelationship to ports formed in the fluid distributor plate, and anotherset of grooves in the other face of said swash plate, identical innumber to that of distributor orifices formed at the end of the barrelcylinders, said bearing further comprising at least three auxiliaryhydrostatic bearings consisting of cavities formed in the other face ofsaid bearing plate, said United States Patent [191 Pruvot 1 HYDROSTATICBEARINGS FOR THE SWASH PLATE OF A BARREL-CYLINDER HYDRAULIC PUMP 0RMOTOR [75] Inventor: Francois C. Pruvot, Billancourt,

France [73] Assignee: Regie Nationale Des Usines Renault, Billancourt,France [22] Filed: Mar. 29, 1972 [21] Appl. No.: 239,290

[30] Foreign Application Priority Data Apr. 28, 1971France.............................. 71.15146 [52] 11.8. 91/489, 91/504[51] Int. FOlb 13/04 [58] Field of Search....................... 91/485,487489 [56] References Cited UNITED STATES PATENTS last-named cavitiesbeing concentric to the grooves of said main bearing.

4 Claims, 16 Drawing Figures Pratt 4/1939 Thoma.. 10/1942 Vickers.....2/1961 Douglas....

3/1961 Schoellhammer HYDROSTATIC BEARINGS FOR THE'SWASH PLATE OF ABARREL-CYLINDER HYDRAULIC PUMP OR MOTOR The present invention relates ingeneral to hydraulic machines such as motors and pumps comprising aplurality of cylinders disposed in a barrel-shape block, wherein thefluid is distributed to the various cylinders through a distributorplate of the flat or revolution type. The term pump will be usedhereinafter to designate a machine according to this invention capableof operating either as a motor or as a pump.

This invention is concerned more particularly with a hydrostatic bearingadapted to balance the impeller plate of a pump of the type mentionedhereinabove and wherein the cylinder or cubic capacity of the pump isvaried by changing the angle measured between the cylinder barrel axisand the pump driving shaft axis, so that a zero angle corresponds to azero capacity.

As a consequence of the specific mode of construction of pumps ofthisgeneral type, the impeller plate rigid with the driving shaft iscaused to support the thrust from the axial pistons. This thrustcomprises a radial component and an axial component, as those skilled inthe art will readily understand.

As a rule, the axial effort is absorbed by one or a plu rality of ballorroller bearings, although pumps types are known wherein plain,hydrodynamic or hydrostatic bearings are used.

More generally, the axial thrust is absorbed by rolling contact thrustbearings whether of the ball-, needleor roller type. In otherconstructions this thrust component is absorbed by hydrostatic bearingssimilar to the bearing described and illustrated in the US. Pat.application Ser. No. 42,71 I filed by the same applicant on June 2, 1970now US. Pat. No. 3,702,576.

In another pump type the swash plate is not driven directly by thedriving shaft and the piston rods are nearly perpendicular thereto. Inthis known arrangement the radial efforts are absorbed by the barrel andthe very moderate radial forces act upon the swash plate, thuspermitting of dispensing with the bearings usually provided forabsorbing these forces.

In the above-mentioned pump types it will be seen that the radial andaxial stresses applied to the swash plate vary as a function of the pumpangle, i.e., the angle formed between the barrel axis and theperpendicular to the swash plate, even if the pressure remains constant.

In pumps comprising means for modifying at will the inclination of theaxis of rotation of the barrel and wherein the swash plate is connecteddirectly to the pump shaft, the axial stress received by the barrel issubstantially constant under constant pressure conditions, irrespectiveof the barrel angle, since said axial stress is subordinate to thenumber and diameter of the pistons.

This made it possible to balance the axial stress received by the barrelby means of simple hydrostatic bearings. As a rule, the axial forcesdeveloping in the cylinders by the working pressure are balanced byproperly shaping the cross-sectional contour of the high-pressure portin the distributor plate. A slight degree of unbalance is neverthelesspreserved as a function of the magnitude of the leakages to betolerated.

However, this solution is ill adapted to the pump swash plate, for theaxial force received thereby varies substantially as the cosine of theangle of inclination of the pump barrel. Now the necessity of reducingthe manufacturing cost, of keeping weight at a low level andperformances at a high level, leads to a substantial, increasingly greatangle of barrel inclination. In certain pumps this angle is at presentas high as 45 corre sponding to an axial thrust exerted against theswash plate which varies in proportion to the cosine of an angle rangingfrom 0 to 45, i.e., a cosine of l to 0.707.

Under these conditions it is clear that the force necessary forhydrostatically compensating the swash plate cannot exceed 0707 timesthe sum of the piston thrusts (representing approximately 65 percent ofthis thrust, since a residual force must constantly be available forkeeping the slide faces in mutual, leak-tight contact). A complementaryaxial bearing should be provided for absorbing an effort correspondingto 35 percent of the piston thrusts when the cylinder capacity of thepump is zero. Of course, these values apply only to a 45 angle of barrelinclination.

To compensate the variation in the axial force during the change in thebarrel inclination angle it is possible for example to modify the feedpressure of the hydrostatic bearing associated with the swash plate bymeans of a valve, so that the variations in the bearing feed pressurecorrespond substantially to the variations in the axial thrust.

Although this compensating device is based on a relatively simpleprinciple, it is attended by several inconveniences, such as:

an increment in the cross-sectional dimensions of the pump, since thepressure modulating valve or valves must be disposed along the pivotaxis of the barrel support;

additional machining operations preventing an economical use of thisbearing type, and

the permanent possibility for the hydrostatic bearing feed pressure tobe either too low or too high, in case frictional contacts develop inanyone of these valves.

If the pressure is too high, it can be reduced by using elements adaptedto produce load losses and disposed in the duct means supplying pressurefluid to the grooves of the hydrostatic bearing, which engage thebearing or sliding surface of the swash plate. These load losses causethe pressure to decrease as the output increases, thus preventing anyswash plate instability. On the other hand, if the feed pressure is toolow, hydrodynamic bearings similar to those equipping the cylinderbarrel are used in order to absorb the non-compensated additionalthrust. This hydrostatic bearing feed pressure variation effectconstitutes in any case an inconvenience.

It is'therefore the primary object of the present invention to provide ahydrostatic bearing adapted to balance the axial thrust of the swashplate in an axial-type, rotary cylinder-block hydraulic pump or motor,whether of the type comprising in-line axial pistons or canting cylinderblocks, with or without connectingrods.

More particularly, this invention is directed to the provision of aneconomical hydrostatic bearing fed from the pressure generated by thepump or the pressure applied to the motor, irrespective of the pressurevalue and the angle of inclination of the swash plate or pump axis inrelation to the axis of rotation of the cylinder barrel.

By way of example the pump proper may be similar (except for thepressure adjustment device) to the pump described in the US. Pat.application Ser. No. 40,541 filed on May 26, 1970 now US. Pat. No.3,661,055. The swash plate of the pump is supported by a radial bearingof which the median plane will be merged if possible into the planecontaining the centres of the ball-joints of the piston connecting-rods,which are mounted in said swash plate.

The pump may also be of a type similar to-the pump type without pistonconnecting-rods, described and illustrated in the U.S. Pat. applicationSer. No. 151,450 filed on June 9, 1971.

The hydrostatic bearing structure for the swash plate of high-pressurecylinder barrel hydraulic machines, which comprises a fluid distributorplate formed with ports communicating with the barrel cylinders, pistonsslidably mounted in said cylinders and connected through ball-joints tosaid swash plate, is characterized in that it is formed essentially onone face of a plate and consists of a main bearing comprising includingon the one hand a first set of two arcuate shallow grooves formed insaid one face registering with a homologue face of the swash plate, saidgrooves being surrounded by sealing surfaces and disposed in properphase relationship to the ports formed in the pump fluid distributorplate, and on the other hand, a second set of grooves formed in saidface of said swash plate, said last-named grooves being shorter thanthose of the first set and equal in number to distributor orificesprovided at the end of the pump barrel cylinder and in proper phaserelationship therewith, said hydrostatic bearing further comprising atleast three elementary auxiliary hydrostatic bearings consisting ofcavities formed in said face of the hydrostatic bearing plate anddisposed concentrically to the grooves of the first set of said mainbearing and on a greater radius than the radius of said grooves of themain bearing.

Other features characterizing this invention will appear as thefollowing description proceeds with reference to the attached drawingillustrating diagrammatically by way of example a balancing orcompensating device according to this invention. In the drawing:

FIG. 1 is a longitudinal view showing a pump equipped with thehydrostatic bearing according to this invention;

FIG. 2 is a plane elevational view showing the hydrostatic bearingdistributor plate, as seen in the direction of the arrows II-I1 of FIG.1;

FIG. 2b is a plane view showing diagrammatically the fluid supply to thehydrostatic bearings of the distributor plate of FIGS. 2 and 3;

FIG. 3 is a plane elevational view of the hydrostatic bearingdistributor plate as seen in the direction of the arrows III-III of FIG.1;

FIG. 4 is a plane elevational view of the swash plate as seen in thedirection of the arrow IV-IV of FIG. 1;

FIG. 5 is a plane view showing a modified form of embodiment of thehydrostatic bearing distributor plate;

FIG. 6 is a plane view showing diagrammatically the fluid supply to thehydrostatic bearings of the distributor plate of FIGS. 2 and 3 in thecase of a pump adopted to operate as a motor;

FIG. 7 is a plane view showing diagrammatically another modified form ofembodiment of the method of supplying fluid to the hydrostatic bearingsof the distributor plate of FIG. 6;

FIG. 8 is a simplified and modified form of embodiment of the plateshown in FIG. 7, in which a duct selector is used;

FIG. 9 is a typical exemplary form of embodiment of the non-returnvalves of FIG. 7;

FIG. 10 is a section taken along the line X-X of FIG. 9; and

FIG. 11 is another modified form of embodiment of the distributor plateof the hydrostatic bearing shown in FIG. 2; and

FIGS. 12 to 15 are sections taken respectively along lines l212, 13-13,l414 and 15-15 of FIG. 2.

Referring first to FIG. 1, the swash plate 1 of a hydraulic barrel pumpis guided radially by a needle bearing 2 and is rotatably assembledthrough splines with a driving shaft 3. A universal joint 4 connectsthis shaft 3 to the cylinder barrel 5 rotatably mounted in antifrictionbearings 6 and inclined to the shaft axis. Pistons 7 are slidablymounted in the cylinders or bores 8 of barrel 5 and operativelyconnected through their rods 9 to the swash plate 1, the end ball-jointsor heads 10 of these connecting-rods being fitted in part-sphericalrecesses 11 formed in one face of swash plate 1.

The opposite face 12 of swash plate 1 bears against a slide face 13 of aplate 14 of a hydrostatic bearing structure according to the presentinvention.

This hydrostatic bearing structure is adapted to withstand the axialstress transmitted to swash plate 1 and adjusts the axial position ofswash plate 1 in order to eliminate frictional contacts between the face13 of bearing plate 14 and the corresponding face 12 of swash plate 1,and also for maintaining a sufficient relative spacing between thesefaces in order to avoid an excessive exposure thereof to the fluidpollution and limit pump leakages to reasonable values. Finally, thebearing must preserve all these properties during variations in thepressure and angle of cant of the pump barrel.

To this end, the swash plate 1 of which one face 12 is illustrated inFIG. 4 comprises a set of arcuate shallow grooves 15 to be subsequentlyreferred to as the second set. The projections of the centres ofballjoints 10 on this face 12 are substantially coincident with thecentres of the corresponding grooves 15 and these grooves are exposed tothe same pressure as the fluid in the corresponding cylinder 8. To thisend, a hole 16 formed in the piston, the passage 17 in connecting-rod 9and another orifice 18 in swash plate 1 are used, this last-namedorifice l8 interconnecting the cavity 11 of ball-joint 10 to the arcuategroove 15.

As shown in FIG. 2, a first set of two arcuate shallow grooves orcavities 19, 20 are formed in the face 13 of bearing plate 14 of thishydrostatic bearing; these grooves or cavities 19, 20 are bound radiallyby arcs having radii R, and R respectively, their lengths being equal tothose of the two radii limiting radially the arcuate grooves 15 of swashplate 1 in FIG. 4.

The arcuate grooves or cavities 19, 20 of the first set have the sameangular arrangement or disposition as the ports 21, 22 of distributorplate 23 contacting the barrel 5 (FIG. 1). However, these grooves areslightly shorter so that the fluid is distributed only through theorifices 24 formed at the end of barrel 5 and registering with ports 21,22 of distributor plate 23. The grooves 19, 20 of bearing plate 14 ofthe hydrostatic bearing are surrounded by sealing surfaces 5,, S Thesame applies to the distributing ports 21, 22 of plate 23.

It is known that as a rule the force pressing the cylinder barrel of apiston pump against its distributor plate is balanced to the extent of90 to 95 percent (although lower and higher values have already beenrecorded) by the reaction forces due to the fluid pressure, saidreaction forces acting upon the barrel in distributing ports and underthe sealing surfaces of said plate.

In the case of the hydrostatic bearing of a swash plate a considerablylower balancing rate is selected. In the present example it iscontemplated to balance 70 per cent of the piston thrust when the barrelaxis has its maximum cant value.

Since the axial force applied to the swash plate varies substantially asthe cosine of the canting angle of the barrel, the force necessary forhydrostatically balancing the swash plate cannot exceed 0.866 times thesum of the piston thrusts, the maximum canting angle of the barrel beingassumed to be equal to 30 (cos 0.866).

The reaction force of grooves 19, and sealing surfaces S and S mentionedhereinabove, with a maximum canting angle of 30, is therefore T 70percent X 0.866, i.e., 60.62 percent of the piston thrust.

In other words, the pressure acting upon the swash plate will balance 70percent of the apparent piston thrust at the maximum canting angle (30).and about 60 percent of the actual piston thrust if the barrel cantingis zero. Under these conditions, auxiliary bearings adapted to absorbthe remaining percentage of the piston thrust must be provided, inaddition to the abovedescribed hydrostatic bearing.

These auxiliary bearings are also hydrostatic bearings and due to therelatively moderate variation in the force to be compensated thethickness of the fluid film between the swash plate and these bearingvaries but very moderately during the variation in the barrel cantingangle.

As shown in FIG. 2, the residual effort of the piston thrust is absorbedby four elementary auxiliary hydrostatic bearings 25, 26, 27 and 28.

Basically, three elementary auxiliary hydrostatic bearings would besufficient, notably if the pump orifices connected to the high-pressureand low-pressure sides are constantly the same. In this case, the threeauxiliary hydrostatic bearings 29, 30 and 31 of a bearing plate 14' aredisposed as shown in FIG. 5. The first and second elementary hydrostaticbearings 29 and 30 are supplied in this case from the high-pressuregroove 20 through an an element (not shown) producing a laminar loss ofpressure. The surface of these two bearings 29, 30 is such that theyabsorb the residual effort under a pressure considerably lower than therated pressure of the pump.

A third auxiliary hydrostatic bearing 31 is also supplied with pressurefluid through the high-pressure groove 20 and its surface is reduced forits load is considerably lower than that applied to the other auxiliaryhydrostatic bearings 29 and 30. It will be noted that the third bearing31 may also be supplied through the lowpressure groove 19, providedhowever that the cramming" pressure of the pump is high enough. In thiscase the surface of the third bearing 31 should be definitely greater.Actually, the arrangement of these three bearings must be such that thepoint of application of the force to be balanced lies within the polygoncontaining the centres of gravity of the surface areas of the auxiliaryhydrostatic bearings.

This arrangement is also applicable to pumps having two directions ofrotation, even if they are operated as motors, provided only that thehigh pressure groove be always the same.

FIG. 6 illustrates diagrammatically a bearing plate 114 of hydrostaticbearing of a swash plate, in the case of a pump adapted to operate as amotor, in either direction of rotation, but wherein the barrel cantingangle may vary from 0 to a maximum value on one side only of the pumpshaft axis.

The auxiliary bearing 125 is supplied with fluid from groove via anelement 132 producing a loss of pressure; this bearing is also connectedto the groove 119 of the main bearing through another element 133 alsoproducing a loss of load. The auxiliary bearing 126 is supplied withfluid from the same groove 119 through an element 134 producing a lossof load; this bearing also communicates with groove 120 through anelement 13S producing a loss of load. Bearing 127 is supplied fromgroove 119 via an element 136 producing a loss of load and bearing 128is supplied from groove 120 via an element 137 also producing a loss ofload.

It is clear that whatever the groove 119 or 120 under pressure, threebearings are constantly supplied directly with high-pressure fluid.

Experience teaches that in the case of a pump comprising seven pistonsit is advantageous that the surface area of each auxiliary hydrostaticbearing 25, 26, 27, 28 lies within the range of 1 to 1.5 times thecrosssectional area of a piston.

The pressure in the various hydrostatic bearings in operation will thenlie in the range of 12 to 21 40 percent of the operating pressure of thepump. The leakage outputs of each hydrostatic bearing are thenrelatively low with respect to the pump output and without anyappreciable influence on its volumetric efficiency. The elementsproducing the aforesaid losses of load are easy to manufacture, as willbe explained presently, without allowing them to become abnormallysensitive to the fluid pollution.

During the variation in the angle of cant of the cylinder barrel 5 thecorresponding variations in the distance between the hydrostaticbearings and the swash plate will be of the order of 0.00l millimeter.

As shown in FIG. 7 illustrating another form of embodiment of fourauxiliary hydrostatic bearings the bearing plate 214 of the hydrostaticbearings of the swash plate comprise two auxiliary bearings 227 and 228fed from grooves 219 and 220, respectively, through elements 236, 237producing the losses of pressure. The other two bearings 225, 226 arefed through the high-pressure groove 219 or 220 through the medium ofcorresponding elements 233, 234 producing losses of pressure, and alsoof check valves 238, 239 adapted to isolate the low-pressure groove.

Referring to FIG. 8, it will be seen that the non-return valves 238, 239of FIG. 7 can be replaced with a fluidway selector 240.

In all the cases contemplated hereinabove the arrangement may be suchthat all the elements producing losses of load be substantially thesam'es in a same pump. This requirement, which is not compulsory, aswill be explained presently, may lead to a substantial simplification inthe manufacturing process. By way of example, the elements producing thelosses of pressure may be constructed by using cylindrical insertscomprising a helical groove.

In the case of mass production it is known that it is advantageous touse sintered metals for making the bearing plates 14 or distributorplates 23, whether on the barrel side or on the swash plate side. Infact, the front face 13 of distributor plate 14 of the hydrostaticbearing may be used as obtained from the sintering pro cess, without anymachining operation, for the balancing rate due to the bearings isrelatively moderate so that the diameters of the auxiliary hydrostaticbearings 25 to 28 or of the sealing surfaces 8,, S need not be machinedwithin extremely accurate tolerances.

As shown on FIGS. 2 and 3 and on the sectional view of FIGS. 12 to 15,orifices 41 and 42 interconnect the grooves 19, of front face 13 of FIG.2 of bearing plate 14 of the hydrostatic bearing to the rear face 43 ofFIG. 3. These orifices open into supply ducts 44 and 45 respectively andthe fluid is directed to elements 46, 47, 48, 49, respectively,consisting of a pair of circular grooves concentric to the axis of theswash plate and formed in the rear face 43 of plate 14 of saidhydrostatic bearing, these elements producing losses pressure along thefluid path as shown in the diagram in FIG. 2b. Orifices 50, 51 connectthe elements 48, 46 aforesaid to the cavities of the auxiliaryhydrostatic bearings 26 and of front face 13. The flow of fluid fromport 19 to the auxiliary fluid bearing 25 passes through orifice 41,duct 44, groove 49 and orifice 51 and to auxiliary bearing 26 throughorifice 41, duct 44, groove 48 and orifice 50. Similarly, the flow offluid from port 20 passes through orifice 42, duct 45, groove 46,orifice 51 to auxiliary bearing 25, and through orifice 42, duct 45,groove 47 and orifice 50 to auxiliary bearing 26.

Similarly, other orifices 52 and 53 connect the grooves 19, 20 of frontface 13 to the rear face 43 of plate 14. The fluid is fed through ducts54, 55 formed in said face 43 and lead into orifices 56, 57 opening inturn into the front face 13.

Duct means 58, 59 formed in the front face 13 and orifices 60, 61 returnthe fluid to the rear face 43; these last-mentioned orifices 60, 61 openat a radius greater than that of groove elements 46, 47, 48 and 49. Theoutlet orifices of the holes formed in the rear face 43 continue assemi-circular grooves 62, 63 concentric to grooves 46 to 49, thusproducing losses of pressure and supplying fluid to orifices 64, 65opening in turn into the cavities of the auxiliary hydrostatic bearings28, 27 formed in the front face 13 of bearing plate 14.

To simplify the drawing all the surfaces of FIGS. 2 and 3 which areflush with the front face 13 or rear face 43 are shown with slighthatches.

A similar procedure could be used for obtaining the elementscorresponding to FIG. 7 which produce the losses of pressure. As shownin FIGS. 9 and 10, the non-return valves 238, 239 may consist forinstance of balls 66 of a diameter slightly inferior to the thickness ofbearing plate 214 and opening on the rear face 243. The cavities 67 andball-valve seats 68 are formed directly in the sintered pieces.

The losses of load or pressure are produced in most instances by grooveshaving width of 0.3 to 1 mm, according to the pump size. These groovesmay be either formed directly by sintering of machined on the rear face43 of plate 14 (or 114, 214). In this last case, the angular extent ofthe grooves producing losses of load are preferably equal or wholemultiples of one another, so that they can be formed simultaneously byusing several tools.

In this case a primary requirement is that the rear face 43 of plate 14be in conformity with the surface to be engaged by this face 43, so thatthe radial leakages are kept within negligible limits. This requirementis easily met by resorting to conventional manufacturing or productionmethods.

The proper operation of the hydrostatic bearing depends mostly on thestability of its bearing plate 14. In fact, if the front face 13 of thisplate receives the total thrust frompistons 7, the various areas of itsrear face 43 will be under pressure. To ensure the stability of plate14, the sum of the forces applied to its rear face must be inferior tothe piston thrust and the moment of the forces exerted on this face mustbe lower than the resultant moment of the piston thrust, so that theplate cannot tilt. The equilibrium of plate 14 is obtained by laterallylimiting the sealing surfaces surrounding the elements producing thelosses of pressure and the fluid inlet ducts.

The hydrostatic bearing according to this invention is advantageouslyapplied to axial pumps of the type described for example in the U.S.Pat. application Ser. No. 151,450 filed on June 9, 1971, alreadymentioned hereinabove. In this pump, the piston shoes engage a rotaryplate of which the lower face bears against a hydrostatic bearing of aswash plate. Since the maximum canting angle of the swash plate of thesepumps is somewhat smaller (about 20) an improved balancing of thegrooves may be expected.

Of course, the specific application described and contemplated hereinshould not be construed as limiting the present invention. Thus, forinstance, the grooves 19, 20 of plate 14 may be supplied either from anexternal source of fluid under pressure, or from the pressure suppliedby the pump itself, but without causing the fluid to flow through thehollow connectingrods. Many other variations, which cannot be describedin detail herein, will readily occur to those skilled in the art withoutdeparting from the scope of the invention.

Other shpaes and arrangements of elements adapted to produce losses ofload may be contemplated and provide substantial advantageous in certaincases.

Thus, for example, if one of the three auxiliary hydrostatic bearingssimultaneously under load is considerably less loaded than the othertwo, it is possible as illustrated in FIG. 11 to dispose between the twogrooves 319, 320 of the main bearings of plate 314 an element 335producing a loss of pressure preventing the less loaded auxiliarybearing (in this case bearing 326 if groove 320 is the high-pressureone) from lifting abnormally the swash plate 1 of the pump. This element335 is connected to the auxiliary bearings 325 and 326 through similarelements 333 and 334, and also through valves 338, 339 to grooves 319,320.

Finally, it will be seen that if groove 320 is the highpressure one thepressure prevailing in auxiliary bearings 325, 326 and 328 respectivelymay be varied by causing a variation, for the same compensation rate, inthe diameters of the edges of groove 320 and in the diameters of thesealing surface.

In fact, the centre of gravity of the forces applied to the swash plateby the pistons is subordinate only to the diameter of the imaginarycylinder containing the axes of the multiple cylinders of the barrel,and also of the cylinder containing the centres of the ball-joints ofthe swash plate; on the other hand, the position of the groove reactiondepends on the groove diameters and also on the diameters of the sealingsurfaces. It will be seen that by varying these diameters it is possibleto shift the point of application of the reaction of the hydrostaticbearings and thus produce a variation in the pressure prevailing in thevarious bearings (the sum of the pressures remaining on the other handconstant).

What is claimed as new is:

l. A hydrostatic bearing structure for the swash plate of ahigh-pressure cylinder-barrel hydraulic machine, which comprises a fluiddistributor plate formed with ports communicating with the barrelcylinders through distribution orifices at the end of said cylinders,and a bearing plate, the bearing structure being formed on only onefront face of the bearing plate and consisting of a main bearingincluding on the one hand a first set of two arcuate shallow groovesformed in said one face registering directly with a homologue rear faceof the swash plate, said grooves being surrounded by sealing surfacesand on the other hand a second set of arcuated grooves formed in saidface of said swash plate, said last-named grooves being shorter thanthose of the first set and equal in number to the distribution orificesprovided at the end of the pump barrel cylinders, said hydrostaticbearing further comprising at least three elementary auxiliaryhydrostatic bearings consisting of cavities formed in said front face ofthe hydrostatic bearing plate and disposed concentrically to the groovesof the first set of said main bearing and on a greater radius than theradius of said grooves of the main bearing, the rear face of saidbearing plate comprising circular recesses concentric to the axis ofsaid plate, said bearing plate having further ducts in its rear facewhich are connected to said recesses of the rear face, a first set oforifices opening through the plate respectively into the grooves of thefront face and into said ducts, and a second set of connecting orificesinterconnecting the recesses on the rear face to said cavities ofelementary auxiliary hydrostatic bearings in the front face.

2. Hydrostatic bearing according to claim 1, wherein at least one ofsaid auxiliary hydrostatic bearings is adapted to be supplied with fluidthrough two separate elements producing losses of pressure and connectedto the grooves of said main hydrostatic bearing, respectively.

3. Hydrostatic bearing according to claim 2, wherein said bearing platecomprises checkvalves interposed between the elements producing a lossof pressure and connected to said auxiliary hydrostatic bearings, andthe grooves of the first set of said main hydrostatic bearing, saidvalves being so disposed as to isolate the low-pressure groove.

4. Hydrostatic bearing according to claim 1, wherein each groove of saidfirst set formed in the front face of said bearing plate delivers fluidunder pressure to a set of at least three auxiliary elementaryhydrostatic bearings through elements producing losses of pressurewhichever the groove of said first set is under pressure, thearrangement of said three auxiliary elementary hydrostatic bearingsbeing such that two bearings are common in said sets of auxiliarybearings and that the point of application of the axial force to bebalanced, which is applied to said swash plate, be constantly locatedwithin the limits of the polygon containing the centers of gravity ofthe surfaces of aforesaid elementary auxiliary hydrostatic bearings.

1. A hydrostatic bearing structure for the swash plate of ahigh-pressure cylinder-barrel hydraulic machine, which comprises a fluiddistributor plate formed with ports communicating with the barrelcylinders through distribution orifices at the end of said cylinders,and a bearing plate, the bearing structure being formed on only onefront face of the bearing plate and consisting of a main bearingincluding on the one hand a first set of two arcuate shallow groovesformed in said one face registering directly with a homologue rear faceof the swash plate, said grooves being surrounded by sealing surfacesand on the other hand a second set of arcuated grooves formed in saidface of said swash plate, said last-named grooves being shorter thanthose of the first set and equal in number to the distribution orificesprovided at the end of the pump barrel cylinders, said hydrostaticbearing further comprising at least three elementary auxiliaryhydrostatic bearings consisting of cavities formed in said front face ofthe hydrostatic bearing plate and disposed concentrically to the groovesof the first set of said main bearing and on a greater radius than theradius of said grooves of the main bearing, the rear face of saidbearing plate comprising circular recesses concentric to the axis ofsaid plate, said bearing plate having further ducts in its rear facewhich are connected to said recesses of the rear face, a first set oforifices opening through the plate respectively into the grooves of thefront face and into said ducts, and a second set of connecting orificesinterconnecting the recesses on the rear face to said cavities ofelementary auxiliary hydrostatic bearings in the front face. 2.Hydrostatic bearing according to claim 1, wherein at least one of saidauxiliary hydrostatic bearings is adapted to be supplied with fluidthrough two separate elements producing losses of pressure and connectedto the grooves of said main hydrostatic bearing, respectively. 3.Hydrostatic bearing according to claim 2, wherein said bearing platecomprises check-valves interposed between the elements producing a lossof pressure and connected to said auxiliary hydrostatic bearings, andthe grooves of the first set of said main hydrostatic bearing, saidvalves being so disposed as to isolate the low-pressure groove. 4.Hydrostatic bearing according to claim 1, wherein each groove of saidfirst set formed in the front face of said bearing plate delivers fluidunder pressure to a set of at least three auxiliary elementaryhydrostatic bearings through elements producing losses of pressurewhichever the groove of said first set is under pressure, thearrangement of said three auxiliary elementary hydrostatic bearingsbeing such that two bearings are common in said sets of auxiliarybearings and that the point of application of the axial force to bebalanced, which is applied to said swash plate, be constantly locatedwithin the limits of the polygon containing the centers of gravity ofthe surfaces of aforesaid elementary auxiliary hydrostatic bearings.